Variable compression ratio mechanism for reciprocating internal combustion engine

ABSTRACT

A variable compression ratio mechanism for a reciprocating internal combustion engine includes upper and lower links mechanically linking a piston pin of a piston to a crankpin, and a control link mechanically linking the lower link to an eccentric cam of a control shaft. Also provided is a control-shaft actuator capable of continuously reducing a compression ratio by driving the control shaft in a first rotational direction and of continuously increasing the compression ratio by driving the control shaft in a second rotational direction opposite to the first rotational direction, so that the compression ratio is controlled to a low value in accordance with an increase in engine speed and/or engine load. A distance from the center of the control shaft to a centerline of the control link, measured with the piston near top dead center, is dimensioned so that the distance continuously decreases as the compression ratio decreases.

TECHNICAL FIELD

The present invention relates to the improvements of a variablecompression ratio mechanism for a reciprocating internal combustionengine.

BACKGROUND ART

In order to vary a compression ratio between the volume existing withinthe engine cylinder with the piston at bottom dead center (BDC) and thevolume in the cylinder with the piston at top dead center (TDC)depending upon engine operating conditions such as engine speed andload, in recent years, there have been proposed and developedmultiple-link type reciprocating piston engines. One such multiple-linktype variable compression ratio mechanism has been disclosed in pages706-711 of the issue for 1997 of the paper “MTZ MotortechnischeZeitschrift 58, No. 11.” The multiple-link type variable compressionratio mechanism disclosed in the paper “MTZ Motortechnische Zeitschrift58, No. 11” is comprised of an upper link mechanically linked at one endto a piston pin, a lower link mechanically linked to both the upper linkand a crankpin of an engine crankshaft, a control shaft arrangedessentially parallel to the axis of the crankshaft and having aneccentric cam whose axis is eccentric to the axis of the control shaft,and a control link rockably or oscillatingly linked at one end onto theeccentric cam of the control shaft and linked at the other end to thelower end of the upper link. By way of rotary motion of the controlshaft, the center of oscillating motion of the control link varies viathe eccentric cam, and thus the distance between the piston pin and thecrankpin also varies. In this manner, a compression ratio can be varied.In the reciprocating engine with such a multiple-link type variablecompression ratio mechanism, the compression ratio is set at arelatively low value at high-load operation to avoid undesired engineknocking from occurring. Conversely, at part-load operation, thecompression ratio is set at a relatively high value to enhance thecombustion efficiency.

SUMMARY OF THE INVENTION

During operation of the reciprocating engine with the multiple-link typevariable compression ratio mechanism, owing to a great piston combustionload (compression pressure) or inertial force a load acts upon theeccentric cam of the control shaft through the piston pin, the upperlink and the control link. That is, owing to the piston combustion load,torque acts to rotate the control shaft in one rotational direction.Assuming that the magnitude of torque occurring due to piston combustionload is excessively great, a driving force needed to drive the controlshaft to a desired angular position and to hold the same at the desiredposition has to be increased. This deteriorates an energy consumptionrate of an energy source such as a motor. In other words, the energysource (i.e., the motor) has to be large-sized. Additionally, in orderto withstand great torque occurring due to piston combustion load, thediameter of the control shaft has to be increased.

Depending on engine/vehicle operating conditions, switching from apart-load operating mode to a highload operating mode frequently occurs.During switching from part-load operation to high-load operation, thecompression ratio is variably controlled to a low compression ratiosuitable to high-load operation. Assuming that switching from high tolow compression ratio is not rapid, engine knocking may occurundesirably. For the above reason, it is desirable to rapidly executeswitching from high to low compression ratio.

Accordingly, it is an object of the invention to provide a variablecompression ratio mechanism for a reciprocating internal combustionengine, which avoids or suppresses the maximum value of torque actingupon a control shaft owing to piston combustion load from excessivelydeveloping during operation of the engine.

It is another object of the invention to enhance the response to switchfrom a control-shaft angular position corresponding to a highcompression ratio suitable for part-load operation to a control-shaftangular position corresponding to a low compression ratio suitable forhigh-load operation in a variable compression ratio mechanism for areciprocating internal combustion engine.

In order to accomplish the aforementioned and other objects of thepresent invention, a variable compression ratio mechanism for areciprocating internal combustion engine comprises a variablecompression ratio mechanism for a reciprocating internal combustionengine including a piston moveable through a stroke in the engine andhaving a piston pin and a crankshaft changing reciprocating motion ofthe piston into rotating motion and having a crankpin, the variablecompression ratio mechanism comprises a plurality of links mechanicallylinking the piston pin to the crankpin, a control shaft extendingparallel to an axis of the crankshaft, an eccentric cam attached to thecontrol shaft so that a center of the eccentric cam is eccentric to acenter of the control shaft, a control link connected at a first end toone of the plurality of links and connected at a second end to theeccentric cam, an actuator that drives the control shaft within apredetermined controlled angular range and holds the control shaft at adesired angular position so that a compression ratio of the enginecontinuously reduces by driving the control shaft in a first rotationaldirection when at least one of engine speed and engine load changes froma first value to a second value higher than the first value and so thatthe compression ratio continuously increases by driving the controlshaft in a second rotational direction opposite to the first rotationaldirection when the at least one of engine speed and engine load changesfrom the second value to the first value, and a distance from the centerof the control shaft to a centerline of the control link passing throughboth a connecting point of the first end and a connecting point of thesecond end, measured with the piston near top dead center, beingdimensioned so that the distance continuously decreases as thecompression ratio decreases.

The other objects and features of this invention will become understoodfrom the following description with reference to the accompanyingdrawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is an assembled view showing a first embodiment of amultiple-link type variable compression ratio mechanism for areciprocating engine, near TDC in a state that the compression ratio iscontrolled to the highest compression ratio.

FIG. 2 is an assembled view showing the multiple-link type variablecompression ratio mechanism of the first embodiment, near TDC in a statethat the compression ratio is controlled to the lowest compressionratio.

FIG. 3 is a predetermined characteristic map showing the relationshipamong engine speed, engine load, and a compression ratio denoted by theGreek letter ε (epsilon).

FIG. 4 shows a characteristic curve illustrating the relationshipbetween a link load F acting upon an eccentric cam of a control shaftthrough a control link (or an engine compression load) and an arm lengthΔD of torque (or an angle α between the centerline of the control linkand the eccentric direction of the center of the eccentric cam to theaxis of the control shaft), in each of the variable compression ratiomechanism of the embodiment and a variable compression ratio mechanismof a comparative example.

FIG. 5 is an enlarged view showing the essential part of the variablecompression ratio mechanism of the first embodiment and used to explainthe operation of the same.

FIG. 6 is an assembled view showing a second embodiment of amultiple-link type variable compression ratio mechanism for areciprocating engine, near TDC in a state that the compression ratio iscontrolled to the highest compression ratio.

FIGS. 7A and 7B respectively show a side view and a cross section of theessential part of a variable compression ratio mechanism of a thirdembodiment.

FIGS. 8A and 8B respectively show a side view and a cross section of theessential part of a variable compression ratio mechanism of a fourthembodiment.

FIGS. 9A and 9B respectively show a side view and a cross section of theessential part of the variable compression ratio mechanism of the firstcomparative example.

FIGS. 10A and 10B respectively show a side view and a cross section ofthe essential part of the variable compression ratio mechanism of thesecond comparative example.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

Referring now to the drawings, particularly to FIG. 1, a cylinder block11 includes engine cylinders 12, each consisting of a cylindrical designfeaturing a smoothly finished inner wall that forms a combustion chamberin combination with a piston 14 and a cylinder head (not shown). A waterjacket 13 is formed in the cylinder block in such a manner as tosurround each engine cylinder. Cylinder 12 serves as a guide forreciprocating motion of piston 14. A piston pin 15 of each of thepistons and a crankpin 17 of an engine crankshaft 16 are mechanicallylinked to each other by means of a multiple-link type variablecompression ratio mechanism (or a multiple-link type piston crankmechanism). In FIGS. 1 and 2, reference sign 18 denotes a counterweight.The linkage of the multiple-link type variable compression ratiomechanism is comprised of three links, namely a lower link 21, arod-shaped upper link 22, and a control link 25. Lower link 21 is fittedonto the outer periphery of crankpin 17 in a manner so as to permitrelative rotation of lower link 21 to crankpin 17. Upper link 22 isprovided to mechanically link the lower link therevia to the piston pin.In order to vary the attitude of each of lower link 21 and upper link22, the variable compression ratio mechanism of the embodiment alsoincludes a control shaft 23 extending parallel to the axis of crankshaft16, that is, arranged in a direction parallel to the cylinder row, andan eccentric cam 24 attached to the control shaft so that the center ofeccentric cam 24 is eccentric to the center of control shaft 23.Eccentric cam 24 and lower link 21 are mechanically linked to each otherthrough control link 25. An actuator 30 (drive means) is provided torotate or drive control shaft 23 within a predetermined controlledangular range and to hold the control shaft at a desired angularposition. The upper end portion of rod-shaped upper link 22 is linked topiston pin 15 in a manner so as to permit relative rotation of upperlink 22 to piston pin 15. The lower end portion of rod-shaped upper link22 is linked or pin-connected to lower link 21 by way of a connectingpin 26, in a manner so as to permit relative rotation of upper link 22to lower link 21. One end (the upper end) of control link 25 is linkedor pin-connected to lower link 21 by way of a connecting pin 27, forrelative rotation. The other end (the lower end) of control link 25 isrotatably fitted onto the outer periphery of eccentric cam 24 forrelative rotation of control link 25 to eccentric cam 24. Actuator 30includes a reciprocating block slider (or a reciprocating piston) 32that reciprocates in an actuator casing 31 and a cylindrical member 34having an internal screw-threaded portion engaged with an externalscrew-threaded portion 33 constructing the rear end portion ofreciprocating block slider 32. In response to a control signal from anelectronic engine control unit often abbreviated to “ECU” (not shown),cylindrical member 34 can be rotated or driven about its axis by meansof a power source such as an electric motor or a hydraulic pump. Thecontrol signal value of the ECU is dependent upon engine operatingconditions such as engine speed and load. Reciprocating block slider 32is arranged in a direction normal to the axis of control shaft 23 insuch a manner as to reciprocate in the actuator casing 31 in the axialdirection of reciprocating block slider 32. A pin 35 is attached to thetip end portion (the front end portion) of reciprocating block slider 32so that the axis of pin 35 is arranged in a direction perpendicular tothe axial direction of reciprocating block slider 32. On the other hand,a control plate 36 is attached to one end of control shaft 23 and has aradially extending slit 37. Pin 35 of reciprocating block slider 32 isslidably fitted into slit 37 of control plate 36.

With the previously-noted arrangement, when cylindrical member 34 isdriven in its one rotational direction in response to a control signalfrom the ECU, one axial sliding movement of reciprocating block slider32, threadably engaged with cylindrical member 34, occurs. Conversely,when cylindrical member 34 is driven in the opposite rotationaldirection in response to a control signal from the ECU, the oppositeaxial sliding movement of reciprocating block slider 32 occurs. In thismanner, the control shaft 23 can be rotated in a desired rotationaldirection based on the control signal from the ECU, with slidingmovement of pin 35 within slit 37. As may be appreciated, actuator 30 isdesigned or constructed so that undesirable reciprocating motion of thereciprocating block slider is prevented by way of engagement between theinternal thread of cylindrical member 34 and the external thread 33 ofreciprocating block slider 32, and so that rotary motion of cylindricalmember 34 is converted into reciprocating motion of reciprocating blockslider 32. In this manner, the center of oscillating motion of controllink 25 fitted onto eccentric cam 24 can be varied by rotating controlshaft 23 depending on engine operating conditions. As a result of this,the attitude of each of upper and lower links 22 and 21 also varies. Acompression ratio of the combustion chamber, that is, a compressionratio between the volume existing within the cylinder with the piston atBDC and the volume in the cylinder with the piston at TDC can bevariably controlled depending upon engine operating conditions. In theshown embodiment, reciprocating block slider 32 moves forwards ordownwards (viewing FIG. 1) and thus control shaft 23 rotates in aclockwise direction ω, the compression ratio can be continuouslyreduced. In contrast, reciprocating block slider 32 moves backwards orupwards (viewing FIG. 1) and thus control shaft 23 rotates in acounterclockwise direction opposite to the direction ω, the compressionratio can be continuously increased.

Referring now to FIG. 3, there is shown the predetermined orpreprogrammed characteristic map showing how the compression ratiodenoted by the Greek letter ε (epsilon) varies relative both enginespeed and engine load. As can be seen from the characteristic map ofFIG. 3, in a high-speed high-load range, the compression ratio is set toa relatively lower value than a low-speed low-load range. In otherwords, in the low-speed low-load range, the compression ratio is set toa relatively higher value than the high-speed high-load range. That is,compression ratio ε is controlled so that compression ratio ε decreasescontinuously as the engine speed increases and so that compression ratioε decreases continuously as the engine load increases.

In the previously-discussed multiple-link type variable compressionratio mechanism of the embodiment, piston pin 15 and crankshaft 16 arelinked to each other through only two links, namely upper and lowerlinks 22 and 21. Therefore, the linkage of the variable compressionratio mechanism of the embodiment is structurally simple. Also, controllink 25 is connected to the lower link instead of connecting to theupper link. Therefore, control link 25 and control shaft 23 can be laidout within a comparatively wide space defined in the lower portion ofthe engine. Thus, it is possible to mount the variable compression ratiomechanism of the embodiment in the engine with comparatively ease.

The multiple-link type variable compression ratio mechanism of the firstembodiment operates as follows. As shown in FIGS. 1 and 2, whencombustion load F1 (the pressure of combustion gas) acts upon the pistoncrown of piston 14 and thus a load F2 is exerted through upper link 22to lower link 21, a link load F is exerted through lower link to controllink 25 so that link load F acts along a control-link centerline L1passing through the axis of connecting pin 27 and the center ofeccentric cam 24. Link load F acts upon eccentric cam 24 via controllink 25, and as a result torque T acts upon control shaft 23 (see FIG.5). Assuming that the distance between the axis of control shaft 23 (orthe center 23 c of control shaft 23) and the center 24 c of eccentriccam 24 is an eccentric distance (simply an eccentricity) H from the axisof control shaft 23 to the center of eccentric cam 24, a line indicativeof the eccentric direction of the center 24 c of eccentric cam 24 to thecenter 23 c of control shaft 23 is denoted by L2, and the angle betweenand the control-link centerline L1 and a line L3 perpendicular to theline L2 is denoted by θ, the aforementioned torque T is derived from theequation T=F·cosθ×H. Additionally, assuming that the distance from thecenter 23 c of control shaft 23 to the control-shaft centerline L1 isdenoted by ΔD, distance ΔD is derived from the equation ΔD=H·cosθ. Thatis, torque T is obtained from the equation T=F·cosθ×H=F·θD. Distance ΔDcorresponds to the arm length of torque T created by link load F. On theassumption that link load F (or combustion load F1) is the same, thelonger the distance ΔD, the greater the torque T. In other words, thelarger the angle α(≦90 degrees) between the control-link center line L1and the line L2 indicative of the eccentric direction of center 24 c ofeccentric cam 24 to center 23 c of control shaft 23, the greater thetorque T. Combustion load F1 (or link load F) becomes maximum with thepiston near or at TDC. Therefore, as appreciated from the characteristiccurve indicated by the solid line in FIG. 4, in the multiple-link typevariable compression ratio mechanism of the first embodiment, distance(arm length) ΔD is dimensioned or set so that distance ΔD continuouslydecreases as link load F increases. That is, distance ΔD continuouslydecreases as compression ratio ε decreases. In other words, angle αbetween the two lines L1 and L2 continuously increases as compressionratio ε increases. By way of proper setting of the distance ΔD, thedistance ΔD (that is, the arm length of torque T created by link load F)tends to reduce when the maximum combustion load F1 (or the maximum linkload F) created at or near TDC increases owing to an increase in engineload or engine speed. Thus, it is possible to suppress thetorque-fluctuation width of torque T fluctuating due to switchingbetween high and low compression ratios. That is to say, duringoperation of the engine, the magnitude of torque T can be leveled orsmoothed. As a result, it is possible to down-size the actuator 30 forcontrol shaft 23. This contributes to down-sizing of the engine itself,improved fuel economy, improved energy efficiency ratio, and down-sizingof control shaft 23. Furthermore, in the variable compression ratiomechanism of the first embodiment, as best seen in FIG. 5, a directionof one force component F_(ω) (equal to F·cosθ and acting in thedirection of line L3) of link load F which load F acts on eccentric cam24 via control link 25 and is created owing to the combustion load at ornear TDC, is set to be the same direction as the rotational direction ωto the low compression ratio. That is, the direction of action of torqueT with the piston at or near TDC is set to be the same direction as therotational direction ω the low compression ratio. When shifting tohighload operation having a possibility of knocking, in other words,when rotating control shaft 23 toward the low compression-ratio side,rotational motion of control shaft 23 toward the low compression-ratioside can be assisted by torque T. This highly enhances the response toswitch from the angular position of control shaft 23 to a control-shaftangular position corresponding to the low compression ratio suitable forthe high-load operation. Therefore, the occurrence of engine knockingcan be certainly prevented, thus enhancing the combustion stability. Inmore detail, in a low-speed low-load range in which the pistoncombustion load Fl is relatively small, there is a tendency for theresponse to switching between low and high compression ratios to belowered. In such a low-speed low-load range, the compression ratio isset to the highest compression ratio (see FIG. 1). Due to setting to thehighest compression ratio, the arm length ΔD of torque T is also set atthe longest distance (substantially corresponding to eccentricity H)near TDC. In other words, the angle α between the two lines L1 and L2 isset at the maximum angle, i.e., substantially 90 degrees near TDC (seeFIG. 4), and therefore the torque value of torque T develops up to themaximum torque level. Owing to the maximum torque value, switching fromhigh to low compression ratio can be smoothly achieved. In contrast tothe above, as appreciated from the characteristic curve indicated by thebroken line in FIG. 4, in the multiple-link type variable compressionratio mechanism of the comparative example, distance (arm length) ΔD isset so that distance ΔD is maximum at the medium compression ratio andrelatively smaller at high and low compression ratios. The arm length ΔDobtained at the high compression ratio is shorter than that obtained atthe medium compression ratio. During the early stages of switching fromhigh to low compression ratio, the switching operation cannot besmoothly achieved, because of the relatively smaller torque Tcorresponding to the high compression ratio. Depending on engine/vehicleoperating conditions, switching of the engine operating mode from thelow-speed low-load range to the medium-speed medium-load rangefrequently occurs. When shifting from the low-speed low-load range tothe medium-speed medium-load range, in other words, when control shaft23 is driven or adjusted from a first angular position corresponding toa high compression ratio to a second angular position corresponding to adesired medium compression ratio, rotary motion of control shaft 23 mustbe stopped rapidly as soon as the control shaft approaches to thedesired medium compression ratio. For this purpose, a counter drivingforce has to be applied to control shaft 23 by means of actuator 30 soas to exert a braking torque to the control shaft. In this case,according to the variable compression ratio mechanism of the embodiment,the arm length ΔD obtained at the medium compression ratio is set to berelatively shorter than that obtained at the high compression ratio. Thetorque T acting on control shaft 23 in the rotational direction ω to thelow compression-ratio side can be properly reduced during shifting fromhigh to medium compression ratio, thus effectively suppressing orreducing the previously-noted counter driving force. This improves theenergy consumption rate. Moreover, In the high-speed high-load range inwhich the magnitude of link load F imparted through control link 25 tocontrol shaft 23 becomes maximum, the engine compression ratio is set atthe lowest compression ratio (see FIG. 3). At the lowest compressionratio, arm length ΔD of torque T becomes the shortest length. As aresult of this, it is possible to effectively properly suppress adriving force that drives or rotates control shaft 23 to the highcompressionratio side against torque T, and/or a holding power thatholds setting of the engine compression ratio to the lowest compressionratio can be effectively suppressed or reduced. It is more preferable toset the distance (arm length) ΔD to substantially “0” near TDC and toset the angle α between L1 and L2 to substantially 0° near TDC, in aparticular condition wherein the engine compression ratio is kept at thelowest compression ratio. In such a case, due to setting to the lowestcompression ratio, torque T can be reduced to as small a torque value aspossible, thus effectively suppressing or reducing a designdriving-force value of driving force produced by actuator 30.

Referring now to FIG. 6, there is shown the cross section of themultiple-link type variable compression ratio mechanism of the secondembodiment. The variable compression ratio mechanism of the secondembodiment of FIG. 6 is similar to the first embodiment of FIGS. 1 and2, except that a line L4 indicative of a longitudinal direction of slit37 of control plate 36 is set to be substantially perpendicular to aline L5 indicative of a direction of reciprocating motion ofreciprocating block slider 32 in the mechanism of the second embodiment.Thus, the same reference signs used to designate elements in themechanism of the first embodiment shown in FIGS. 1 and 2 will be appliedto the corresponding reference signs used in the mechanism of the secondembodiment shown in FIG. 6, for the purpose of comparison of the firstand second embodiments. Detailed description of the same elements willbe omitted because the above description thereon seems to beself-explanatory. In case of the perpendicular layout between line L4indicative of the longitudinal direction of slit 37 of control plate 36and line L5 indicative of the direction of reciprocating motion ofreciprocating block slider 32, a direction of action of a load exertedfrom control shaft 23 to reciprocating block slider 32 near TDC owing tothe piston combustion load is set to be the same direction as thedirection of reciprocating motion of reciprocating block slider 32, withthe compression ratio set at the highest compression ratio at which thepossibility of knocking is high and thus a higher response to switchingfrom high to low compression ratio is required. As a consequence, aninstantaneous speed reduction ratio or an instantaneous decelerationrate of a power-transmission mechanism that transmits from a powersource such as an electric motor or a hydraulic pump to control shaft 23can be effectively reduced. Owing to the reduced instantaneous reductionratio arising from the previously-noted perpendicular layout, theswitching operation from high to low compression ratio can beeffectively assisted by virtue of piston combustion load F1. Thus, it ispossible to remarkably enhance the response to switching ofreciprocating block slider 32 to the low compression-ratio side.

Good and poor lubricating-oil passage layouts are explained hereunder inreference to FIGS. 7A through 10B. FIGS. 7A and 7B show the goodlubricating-oil passage layout used in the variable compression ratiomechanism of the third embodiment. FIGS. 8A and 8B show the goodlubricating-oil passage layout used in the variable compression ratiomechanism of the fourth embodiment. On the other hand, FIGS. 9A and 9Bshow the poor lubricating-oil passage layout used in the variablecompression ratio mechanism of the first comparative example. FIGS. 10Aand 10B show the poor lubricating-oil passage layout used in thevariable compression ratio mechanism of the second comparative example.

As shown in FIGS. 7A-10B, the control shaft 23 (including eccentric cam24) is formed therein with first and second lubricating-oil passageportions 40 and 41, in order to feed lubrication oil to the shaftjournal portion of control shaft 23. First lubricating-oil passageportion 40 is axially formed in the control shaft in a manner so as topass the interior of control shaft 23 and the interior of eccentric cam24 and to axially extend parallel to the axis of control shaft 23. Onthe other hand, second lubricating-oil passage portion 41 is a straightoil passage formed in the eccentric cam in a manner so as to pass theinterior of eccentric cam 24 and to extend in a direction perpendicularto the axially-extending first lubricating-oil passage portion 40. Aninlet port 42 of second oil-lubricating passage portion 41 is opened tofirst oil-lubricating passage portion 40. On the other hand, an outletport 43 of second oil-lubricating passage portion 41 is opened into aclearance space defined between the bearing surface 25 a of control link25 and the outer peripheral surface 24 a of eccentric cam 24. Outerperipheral surface 24 a is opposite to and in sliding-contact withbearing surface 25 a. As shown in FIGS. 9A, 9B, 10A and 10B, if outletport 43 of second oil-lubricating passage portion 41 is laid out in thevicinity of control-link centerline L1 near TDC in a state where thecompression ratio is set at the lowest compression ratio, there are somedrawbacks. For example, as shown in FIGS. 9A and 9B, when outlet port 43is laid out along control-link centerline L1 in a side (the upper side)opposite to the axis of control shaft 23, lubricating oil is fed intothe widest space (the maximum bearing clearance space) defined betweenthe two opposing surfaces 25 a and 24 a. Most of lubricating oil fedinto the clearance is wastefully flown out in the cross direction of theshaft journal portion of eccentric cam 24. In contrast, as shown inFIGS. 10A and 10B, when outlet port 43 is laid out along control-linkcenterline L1 in the other side (the lower side) facing the axis ofcontrol shaft 23, outlet port 43 is located in the high-bearing-pressurearea of maximum loading. In such a case, the effectivepressure-receiving area of the shaft bearing portion may be reducedundesirably. As set out above, in the case that outlet port 43 is laidout to be in alignment with control-link centerline L1 and its vicinitywith the piston near TDC in a state where the compression ratio is setto the lowest compression ratio, sufficient lubricating effect cannot beprovided.

From the viewpoint as discussed above, in the variable compression ratiomechanism of each of the third (FIGS. 7A and 7B) and fourth (FIGS. 8Aand 8B) embodiments, as viewed from the lateral cross section shown inFIG. 7B or 8B, outlet port 43 of second oil-lubricating passage portion41 is laid out in such a manner as to be spaced apart from each of twointersection points of the circumference of eccentric cam 24 andcontrol-link centerline L1 or apart from the vicinity of each of the twointersection points. Concretely, outlet port 43 is laid out at or nearbya position of outer peripheral surface 24 a of eccentric cam 24 thatcrosses a line passing through eccentric-cam center 24 c and arrangedperpendicular to control-link centerline L1, so that the distance fromoutlet port 43 to control-link centerline L1 is substantially maximum.In the third embodiment shown in FIGS. 7A and 7B, only one secondlubricating-oil passage portion 41 is formed in each of eccentric cams24 and therefore outlet port 43 is arranged on one side of control-linkcenterline L1. In the fourth embodiment shown in FIGS. 8A and 8B, twosecond lubricating-oil passage portions (41, 41) are formed in each ofeccentric cams 24 and therefore two outlet ports (43, 43) arerespectively arranged on both sides of control-link centerline L1 sothat these outlet ports (43, 43) are diametrically opposed to each otherwith respect to the center (or axis) of eccentric cam 24. Owing to thegood lubricating-oil passage layout, in particular owing to the goodlayout of outlet port 43 of second lubricating-oil passage portion 41,it is possible to provide sufficient lubrication of the shaft journalportion of eccentric cam 24 and sufficient lubrication of the bearingportion of control link 25 by way of lubricating oil supplied ordischarged into the middle-pressure area through outlet port 43 ofsecond lubricating-oil passage portion 41, without lowering thepressure-receiving surface.

The entire contents of Japanese Patent Application No. P2000-311562(filed Oct. 12, 2000) is incorporated herein by reference.

While the foregoing is a description of the preferred embodimentscarried out the invention, it will be understood that the invention isnot limited to the particular embodiments shown and described herein,but that various changes and modifications may be made without departingfrom the scope or spirit of this invention as defined by the followingclaims.

What is claimed is:
 1. A variable compression ratio mechanism for areciprocating internal combustion engine including a piston moveablethrough a stroke in the engine and having a piston pin and a crankshaftchanging reciprocating motion of the piston into rotating motion andhaving a crankpin, the variable compression ratio mechanism comprising:a plurality of links mechanically linking the piston pin to thecrankpin; a control shaft extending parallel to an axis of thecrankshaft; an eccentric cam attached to the control shaft so that acenter of the eccentric cam is eccentric to a center of the controlshaft; a control link connected at a first end to one of the pluralityof links and connected at a second end to the eccentric cam; an actuatorthat drives the control shaft within a predetermined controlled angularrange and holds the control shaft at a desired angular position so thata compression ratio of the engine continuously reduces by driving thecontrol shaft in a first rotational direction when at least one ofengine speed and engine load changes from a first value to a secondvalue higher than the first value and so that the compression ratiocontinuously increases by driving the control shaft in a secondrotational direction opposite to the first rotational direction when theat least one of engine speed and engine load changes from the secondvalue to the first value; and a distance from the center of the controlshaft to a centerline of the control link passing through bothconnecting points of the first and second ends, measured with the pistonnear top dead center, being dimensioned so that the distancecontinuously decreases as the compression ratio decreases.
 2. Thevariable compression ratio mechanism as claimed in claim 1, wherein adirection of one force component of a load acting on the eccentric camvia the control link owing to combustion load acting on the piston nearthe top dead center, is set to be identical to the first rotationaldirection, the one force component acting in a direction of a lineperpendicular to a line indicative of an eccentric direction of thecenter of the eccentric cam to the center of the control shaft.
 3. Thevariable compression ratio mechanism as claimed in claim 2, wherein anangle between the centerline of the control link and the line indicativeof the eccentric direction is set to be substantially 90 degrees withthe piston near the top dead center in a state where the compressionratio is set at a highest compression ratio.
 4. The variable compressionratio mechanism as claimed in claim 2, wherein the distance from thecenter of the control shaft to the centerline of the control link is setto be substantially 0 with the piston near the top dead center in astate where the compression ratio is set at a lowest compression ratio.5. The variable compression ratio mechanism as claimed in claim 1,wherein the plurality of links comprises an upper link connected at oneend to the piston pin and a lower link connected to both the crankpinand the other end of the upper link, and one end of the control shaft isconnected to the lower link through the control link.
 6. The variablecompression ratio mechanism as claimed in claim 1, wherein the actuatorcomprises a reciprocating block slider capable of reciprocating in adirection normal to an axis of the control shaft, and the reciprocatingblock slider has a pin attached to a tip end portion of thereciprocating block slider and the control shaft has aradially-extending slit formed at its shaft end, and a line indicativeof a longitudinal direction of the slit is set to be substantiallyperpendicular to a line indicative of a direction of reciprocatingmotion of the reciprocating block slider in a state where thecompression ratio is set at a highest compression ratio.
 7. The variablecompression ratio mechanism as claimed in claim 1, wherein the controlshaft and the eccentric cam have a lubricating-oil passage formedtherein, and an outlet port of the lubricating-oil passage is openedinto a clearance space defined between a bearing surface of the controllink and an outer peripheral surface of the eccentric cam being insliding-contact with the bearing surface of the control link, and theoutlet port is laid out to be out of alignment with the centerline ofthe control link and its vicinity with the piston near the top deadcenter in a state where the compression ratio is set at a lowestcompression ratio.
 8. The variable compression ratio mechanism asclaimed in claim 7, wherein the lubricating-oil passage comprises afirst lubricating-oil passage portion formed in the control shaft andextending parallel to the axis of the control shaft and a secondlubricating-oil passage portion formed in the eccentric cam andextending in a direction perpendicular to the first lubricating-oilpassage portion, and an inlet port of the second lubricating-oil passageportion is opened to the first lubricating-oil passage portion while anoutlet port of the second lubricating-oil passage portion is opened intothe clearance space defined between the bearing surface of the controllink and the outer peripheral surface of the eccentric cam.
 9. Thevariable compression ratio mechanism as claimed in claim 8, wherein theoutlet port is laid out at or nearby a position of the outer peripheralsurface of the eccentric cam that crosses a line passing through thecenter of the eccentric cam and arranged perpendicular to the centerlineof the control link so that a distance from the outlet port to thecenterline of the control link is substantially maximum, with the pistonnear the top dead center in the state where the compression ratio is setat the lowest compression ratio.
 10. The variable compression ratiomechanism as claimed in claim 9, wherein two second lubricating-oilpassage portions are formed in the eccentric cam and two outlet portsare respectively arranged on both sides of the centerline of the controllink so that the two outlet ports are diametrically opposed to eachother with respect to the center of the eccentric cam.